Rotary actuator with shortest force path configuration

ABSTRACT

A parallel eccentric electro-mechanical actuator provides motive power and includes an electric prime mover that drives the reducer&#39;s pinion gear. This pinion drives minimum three star gears with stationary shaft bearings. Each shaft contains an eccentric which are completely in parallel with each other. These eccentrics can be thought of as parallel/in-phase driven crankshafts. Each eccentric drives the parallel eccentric (PE) gear through a bearing. The PE gear exhibits a circular motion (without rotation) which in itself is unbalanced. The crankshafts have another eccentric which create an opposite inertia force to balance that of the PE gear. The PE gear contains an external toothed gear on its periphery. It meshes with the internal teeth of the output ring gear. The relative motion between the PE gear and the ring gear is that the PE gear rolls inside the ring gear. This relative motion is called hypo-cycloidal motion.

RELATED APPLICATIONS

This application is a continuation-in-part application of U.S. patentapplication Ser. No. 13/593,397, filed Aug. 23, 2012, now pending,having the same title and the same inventor, and which is incorporatedherein by reference in its entirety; which application is a continuationapplication of U.S. patent application Ser. No. 13/236,767, filed Sep.20, 2011, now abandoned, and which is incorporated herein by referencein its entirety; which application is a divisional application of U.S.patent application Ser. No. 12/032,931, filed Feb. 18, 2008, now U.S.Pat. No. 8,033,942, issued Oct. 11, 2011, and which is incorporatedherein by reference in its entirety; which application is acontinuation-in-part application of U.S. patent application Ser. No.11/288,042, filed Nov. 28, 2005, now U.S. Pat. No. 7,431,676, issuedOct. 7, 2008, and which is incorporated herein by reference in itsentirety; which application is a continuation-in-part application ofU.S. patent application Ser. No. 10/714,183, filed Nov. 14, 2003, nowU.S. Pat. No. 7,081,062, issued Jul. 25, 2006, and which is incorporatedherein by reference in its entirety; which application claims thebenefit of U.S. provisional application No. 60/429,276, filed Nov. 25,2002, and which is incorporated herein by reference in its entirety; andthe benefit of U.S. Provisional application No. 60/890,246, filed Feb.16, 2007, which is incorporated herein by reference in its entirety.

FIELD OF THE DISCLOSURE

The present disclosure relates generally to mechanical actuators, andmore particularly to rotary actuators having a shortest force pathconfiguration.

DESCRIPTION OF THE RELATED ART

Currently the design of mechanical systems using actuators is a complextask with many options to choose from, each option with inherent flawsand shortcomings. The narrow reduction range of these options preventsthe system from being versatile, which may be expensive and wasteful.Customizing the system to such small ranges creates specializedequipment. Specialized equipment is expensive to build, maintain, andoperate.

At this time there is no proven ideal reducer between 25-to-1 and75-to-1. From 25:45-to-1, the two stage star compound does very well.From 35-75-to-1, the differencing epicyclic with three or more planetsdoes quite well. However, if an application requires reduction over thewhole range, multiple units are going to be needed.

Two parallel 4-bar linkages (equal parallel cranks) when hooked together(90° from each other at the cranks) provide pure circular motion (theradius of the cranks) for all points in the connecting rod even thoughthe connecting rod itself does not rotate. Should these cranks becomevery small, they can be physically embodied as eccentrics. This couldproperly be called parallel eccentric motion which will be the basis forthe gear reducer developed here. Note that two cranks at 180° areindeterminate in the central position for a single input driver.Usually, it is best to use three or more such cranks (eccentrics) drivensimultaneously through a star compound gear arrangement to provide thesmoothest continuous motion.

Currently the best practice commercial electro-mechanical actuator isthe Nabtesco gear train. Note that the Nabtesco uses cycloidal motionwith cycloidal wave teeth. The Nabtesco suffers from too high a loadcarrying demand on its crankshaft bearings, long force path, a highpressure angle, significantly higher inherent mesh sliding friction, anda damaging (and unfixable) lost motion.

Developing a simple high load capacity electro-mechanical actuator tobest meet a larger reduction range will solve many of these problems.The unit should engage the maximum number of teeth in the mesh, exhibitthe least lost motion, and contain a nominal level of effective inertia.

SUMMARY

As discussed above, there is a diverse environment of electro-mechanicalactuators capable of specific reduction ratios. The inventor of theparallel eccentric electro-mechanical actuator described herein believesthat the effectiveness of a single actuator that covers a larger part ofthe reduction range would drastically improve the designs andversatility of the larger systems. As well, it is believed that morerobust and cost effective designs can be produced with an actuatorcapable of spanning the larger reduction range.

According to one aspect of the present disclosure, the disclosedparallel eccentric electro-mechanical actuator provides motive power andincludes an electric prime mover that drives the pinion gear of thereducer. This pinion drives three (or more) star gears with stationaryshaft bearings. Each shaft contains an eccentric e (all the same size)which are completely in parallel with each other (in phase). Theseeccentrics can be thought of as three parallel/in-phase drivencrankshafts. Each eccentric drives the parallel eccentric (PE) gearthrough a bearing. The PE gear exhibits a circular motion (withoutrotation) which in itself is unbalanced. To counteract this imbalance,the crankshafts have another eccentric with magnitude k×e which drives aPE balance mass. This PE mass must create an opposite inertia force tobalance that of the PE gear (which includes the other unbalanced massesassociated with the eccentrics and crankshaft bearings). Hence, theweight of the PE mass can be reduced according to the magnitude of theeccentric ratio k. The PE gear contains an external toothed gear on itsperiphery. It meshes with the internal teeth of the output ring gear.The relative motion between the PE gear and the ring gear is that the PEgear rolls inside the ring gear. This relative motion is calledhypo-cycloidal motion.

According to another aspect of the disclosure, a rotary actuator isdisclosed whose construction is based on shortest force path principles.In a preferred embodiment, the rotary actuator comprises an outputattachment plate, an outer attachment shell and a principal bearing, andthe principal bearing has a first surface which is attached to theoutput attachment plate, and a second surface which is attached to theouter attachment shell. In an especially preferred embodiment, theoutput attachment plate, outer attachment shell and a principal bearingare arranged such that a line, cylinder or annulus exists which passesthrough the output attachment plate, the principal bearing and the outerattachment shell.

These and other advantages of the disclosed subject matter, as well asadditional novel features, will be apparent from the descriptionprovided herein. The intent of this summary is not to be a comprehensivedescription of the claimed subject matter, but rather to provide a shortoverview of some of the subject matter's functionality. Other systems,methods, features, and advantages here provided will become apparent toone with skill in the art upon examination of the following FIGUREs anddetailed description. It is intended that all such additional systems,methods feature, and advantages be included within this description, bewithin the scope of the accompanying claims.

BRIEF DESCRIPTION OF THE DRAWINGS

For a more complete understanding of the features and advantages of thepresent disclosure, reference is now made to the detailed description ofthe disclosure along with the accompanying figures in whichcorresponding numerals in the different figures refer to correspondingparts and in which:

FIG. 1 is a cutaway isometric view of a rotary actuator in accordancewith one embodiment of the present disclosure;

FIG. 2 is a cutaway isometric view of a rotary actuator in accordancewith a second embodiment of the present disclosure;

FIG. 3 is a cutaway isometric view of a rotary actuator in accordancewith a third embodiment of the present disclosure;

FIG. 4 is a cutaway isometric view of a rotary actuator in accordancewith a fourth embodiment of the present disclosure;

FIG. 5 is a cutaway isometric view of a rotary actuator in accordancewith a fifth embodiment of the present disclosure;

FIG. 6 is a cutaway isometric view of a rotary actuator in accordancewith a sixth embodiment of the present disclosure;

FIG. 7 is a cutaway isometric view of a rotary actuator in accordancewith a seventh embodiment of the present disclosure;

FIG. 8 is a cutaway isometric view of a rotary actuator in accordancewith certain embodiments of the present disclosure;

FIG. 9 is a side view of a circular arc gear tooth mesh in accordancewith certain embodiments of the present disclosure;

FIG. 10 is a side view of a single circular arc gear tooth in accordancewith certain embodiments of the present disclosure;

FIG. 11 is a side view of a single circular arc gear tooth in accordancewith certain embodiments of the present disclosure;

FIG. 12 is a side view of a single circular arc gear tooth in accordancewith certain embodiments of the present disclosure;

FIGS. 13 through 18 present a variety of exemplary systems in which thedisclosed subject matter may find beneficial application;

FIG. 19 is a cross-section view of a parallel eccentricelectro-mechanical actuator in accordance with one embodiment of thepresent disclosure;

FIG. 20 is a cutaway isometric view of a parallel eccentricelectro-mechanical actuator in accordance with one embodiment of thepresent disclosure;

FIG. 21 is a cross-section view of a parallel eccentricelectro-mechanical actuator in accordance with a second embodiment ofthe present disclosure;

FIG. 22 is a cutaway isometric view of a parallel eccentricelectro-mechanical actuator in accordance with a second embodiment ofthe present disclosure;

FIG. 23 is a cross-section view of a parallel eccentricelectro-mechanical actuator in accordance with a third embodiment of thepresent disclosure;

FIG. 24 is a cutaway isometric view of a parallel eccentricelectro-mechanical actuator in accordance with a third embodiment of thepresent disclosure;

FIG. 25 is a cross-section view of a parallel eccentricelectro-mechanical actuator in accordance with a fourth embodiment ofthe present disclosure; and

FIG. 26 is a cutaway isometric view of a parallel eccentricelectro-mechanical actuator in accordance with a fourth embodiment ofthe present disclosure.

FIG. 27 is a cutaway isometric view of the rotary actuator of FIG. 2which has been annotated to depict the annular shortest force path thatexists in this actuator.

DETAILED DESCRIPTION

While the making and using of various embodiments of the presentdisclosure are discussed in detail below, it should be appreciated thatthe present disclosure provides many applicable inventive concepts,which can be embodied in a wide variety of specific contexts. Thespecific embodiments discussed herein are merely illustrative ofspecific ways to make and use the disclosure and do not delimit thescope of the disclosure.

Certain embodiments of the present disclosure are standardized rotaryactuators which can be mass produced at low cost and still maintain ahigh level of performance. Various levels of ruggedness—for example,light, medium, and heavy—may be employed for various applications. Infact, certain of the actuator concepts described herein will be foundsufficiently rugged in their basic design that parts made of plastic orformed metal can be used to reduce cost while still providing ahighly-versatile actuator useful in a variety of applications. Theseapplications may include, but are not limited to, portable tools,educational robots, toys, and automobiles.

The present disclosure is a new high-performance rotary actuator in avariety of embodiments sharing certain characteristic features.Depending on the application, each of the embodiments incorporatesfeatures generating one or more of certain advantages.

The novel design of certain embodiments of the present disclosureprovide simplicity of design using a relatively small number of partsand a minimized list of parameters, thereby allowing for a relativelysmall form factor exhibiting exceptional compactness, stiffness and loadcapacity, along with quiet and efficient operation. They are designedfor easy assembly, ideal for mass production at various quality levels.

In general, these actuators are of exceptionally rugged design,exhibiting relative insensitivity to temperature and tolerance effects.The actuators of the present disclosure incorporate a relatively shortforce path across a high-stiffness cross-roller bearing, therebymaximizing stiffness and strength. In certain embodiments, the actuatorsof the present disclosure incorporate standardized attachmentarchitectures.

In order to promote standardization, the rotary actuator of the presentdisclosure can be configured to serve as a standardized “building block”within a system. Such a building block may, for example, be intelligentand adaptable, provide for a maximum performance envelope, be compactand rugged, be optimized in its structural design, provide standardizedinterfaces for quick replacement by technicians anywhere in the world,and be produced in large enough quantities to take advantage ofeconomies of scale in manufacturing.

Prime mover requirements may be met either by D.C. brushless motors orswitched reluctance motors, either in cylindrical or pancake format. Thegear trains may be made unique, compact, rugged and cost effective underproduction in large quantities.

In certain embodiments, cross-roller bearings are used to form the jointbearings themselves. Cross-roller bearings are selected not only fortheir stiffness but also owing to their proper geometric configuration.In certain embodiments, the cross-roller bearing or similar largediameter bearing acts as the principal gear train bearing at the sametime. Precision large and small-scale actuators can be used separatelyor combined to satisfy demanding positional accuracy requirements.

In manufacturing cells, the rotary actuator modules of the presentdisclosure may be used directly as simple transfer devices, drivers ofconveyers, or joint actuators in 2 degree-of-freedom manipulators. Atthe other end of the complexity continuum, highly dexterous manipulatorshaving 10 degrees of freedom and above can be assembled on demand. Eachof the above systems could be assembled as needed, all with the sameinterfaces, and all with the same maintenance requirements, perhaps fromonly 5 basic sizes in each cell application, and all driven by oneuniversal software package to reduce cost, increase performance, and toopen up the architecture of such systems.

FIG. 1 depicts an isometric cutaway view of a rotary actuator 10 inaccordance with one embodiment of the present disclosure. Rotaryactuator 10 shown in FIG. 1 may be configured to be very rugged, havinghigh levels of both stiffness and shock resistance.

A rotary actuator such as rotary actuator 10 can have a number ofgeometrical configurations. In one such configuration, a rotary actuatorhas a “pancake” geometry, being relatively narrow in thickness along itscenterline and relatively large in diameter. Rotary actuator 10 shown inFIG. 1 has such geometry. In alternate embodiments, a rotary actuatormay have a “coffee can” geometry, being relatively wide along thecenterline and relatively small in diameter. Rotary actuator 50 shown inFIG. 2 has this type of geometry.

Generally, the pancake version is driven by a switched reluctance motor(SRM) and optimized to produce higher torques at lower speeds. The“coffee can” version is generally optimized for use in slim/dexterousmachines such as serial robot manipulators. This version is usuallydriven by a brushless D.C. motor of somewhat lower torque and higherspeed ranges as compared to the SRM.

In general, it is desirable to satisfy as many design objectives aspossible while at the same time minimizing complexity. This combinationof design criteria argue in favor of combining functions when possible.In certain embodiments, the rotary actuator of the present disclosure isconstructed so as to not merely provide rotary power to a joint, but tofunction as the joint itself, incorporating sufficient structuralrigidity as to make additional rotary bearing structure extraneous.

In certain embodiments, the incorporation of quick-change interfacesinto the input/output attachment structures of the rotary actuatorsprovides the designer with the ability to assemble machines on demand.The geometry of one embodiment of such a quick-change interface isdescribed in detail in connection with FIG. 8, below. In certainembodiments, the self-contained actuator may incorporate 80% or more ofthe machine's complexity, including electronics, brakes, buses, sensors,bearings, motor, gear train, and all necessary attachments andinterfaces.

Rotary actuator 10 makes use of internal bull gear 14 and sun gear 16 aspart of the attachment components of the rotary actuator 10, separatedby a principal cross-roller bearing 18. The bull gear 14 and sun gear 16are driven by planet gears 20 and 22 supported by bearings 24 on pressfit shafts 26 passing through the sides of the planet cage 28.

Because the bull gear 14 and sun gear 16 are part of the structure ofthe rotary actuator 10, the required weight goes down while thestiffness goes up. Also, because this design employs a large diametercross-roller bearing 18, the structural stiffness of the rotary actuator10 is also greatly improved. In certain embodiments, the bearing racescan be machined directly into the bull gear 14 and/or sun gear 16 so asto improve the structural integrity of the design. A ball bearing may beused in place of cross-roller bearing 18 in less-demanding applications.Accordingly, the structure of the rotary actuator 10 can be made muchsmaller, lighter in weight, and more cost effective, through a reductionin the number of parts and simplified assembly. The planet gears 20 and22 may be used in a Ferguson paradox gear train mechanism to furtherimprove manufacturing simplicity.

Magnet disk 30 of the prime mover 32 is rigidly attached to the planetcage 28 to form the simplest possible configuration between the primemover 32 and the gear train 34. Planet cage 28 and magnet disk 30 aresupported by bearing 36 in the bull gear 14 and needle bearings 38 onstationary shaft 40. This design provides a very rugged support for themoving structure of rotary actuator 10 so as to best resist shock.

Planet cage 28 can be made lighter in order to reduce inertia in caseswhere additional responsiveness is desirable. The number of planets 20and 22 may be as small as 2 or as large as 9 depending on the relativedimensions, speed, desired stiffness, inertia requirement, tooth sizingrequired for loading, and other factors.

Bearing 42 on shaft 40 is used to provide additional support to theoutput attachment plate 44 of rotary actuator 10. Where stiffness is animportant consideration, the attachments to the neighboring structureson shell 12 and plate 44 may be placed in close proximity tocross-roller bearing 18 in order to maximize the resulting structuralstiffness of the system. In rotary actuator 10, field 46 is larger thanmagnet disk 30. This additional size accommodates end turns in the field46.

The switched reluctance motor (SRM) geometry shown in FIG. 1 is designedto maximize torque, and this design may be optimized for applicationswherein high rotational speed is not a principal concern. A wide varietyof aspect ratio considerations may be met employing both the SRM and DCprime movers. Given a cylindrical prime mover such as a D.C. brushlessmotor of higher relative speed and lower relative torque as compared tothe switched reluctance motor, the geometry of rotary actuator 10 can bemodified into a coffee can geometry having all the other attributes ofthe pancake-shaped rotary actuator 10. Such an actuator is shown in FIG.2 and generally designated 50.

In one embodiment, rotary actuator 50 may operate at speeds as much asten times higher, but produce ten times less torque, than rotaryactuator 10 of FIG. 1. In rotary actuator 50, there is a much higherconcern for inertia in the moving structure and less concern for stressin the gear teeth due to a lower expected torque capacity.

FIG. 2 depicts rotary actuator 50 in an isometric cutaway view inaccordance with a second embodiment of the present disclosure. Rotaryactuator 50 is typically longer than rotary actuator 10, and there ismore concern for the stiffness of the planet cage 68. Accordingly,additional support is provided by bearing 76, embedded in the stiffattachment shell 52 of actuator 50. In order to simplify the design ofrotary actuator 50, the planets 60 and 62 are supported by bearings 64which ride on shafts 66, which are press fit into the planet cage 68 tofurther increase the stiffness of planet cage 68.

The output attachment plate 84 and central stationary shaft 230 aremutually supported by bearing 82. Generally, because of highervelocities in the D.C. motor, the structure of the planet cage 68 willbe lightened to reduce inertia and the bearings 58, 64, and 76 will bechosen for this higher velocity regime.

As will be appreciated by those of skill in the art, additional planetstend to increase stiffness, reduce backlash, and improve positionalaccuracy at the expense of complexity and increased inertia. Large geartrain ratios require the use of multiple stages or Ferguson Paradox typeepicyclic gear trains. Generally, the planet gear cage will representthe most complex part of the rotary actuator, adding to cost,complexity, and assembly issues.

In alternate embodiments, compound gears can be used in certain cases.Such gear trains incorporate, however, inherent limitations. These typesof gear trains can give a realistic reduction of no greater than10-to-1. Further, these gear trains tend to exhibit considerablebacklash and have high rotary inertia. Finally, they are insufficientlyrigid in rotary compliance, are heavy and are not space efficient.

These significant limitations may be substantially reduced or eliminatedby arranging multiples of a second gear of a compound gear train in asymmetric array about a first gear for forming a “star compound” geartrain. The star compound gear train provides multiple meshes with theinput pinion (the first gear), results in no unbalanced forces on thepinion, provides for four to six teeth in mesh, reduces contact andbending stresses, and provides for a compact concentric configuration tomatch the concentric geometry of the prime mover.

Accordingly, epicyclic gear trains are better for rotary actuatorsbecause of their compatible geometry to the rotary prime mover.Unfortunately, these gear trains exhibit limitations as well. Themaximum realistic gear reduction of such a mechanism is on the order of100:1. Compound epicycle gear trains can, of course, provide reductionshigher than 100:1 through the use of multiple stages. Compound geartrains, however, incorporate the limitations described above. Ingeneral, epicyclic gear trains exhibit a significant degree of backlash,require high tolerances, and are temperature sensitive. In fact,backlash generally must be designed in to account fortemperature-related dimensional changes. Finally, the involute gearteeth used in epicyclic gear trains are often designed to be relativelytall, in order to maintain between one to two teeth in mesh. Thisgeometry increases the loading at the root as well as sliding velocity,reducing both the strength and the efficiency of the mechanism.

In order to overcome the above limitations of epicyclic gear trains,elements are described below employing a single planet driven by aneccentric to make a “wobble” plate design while satisfying all thekinematic requirements normally associated with epicyclic gear trains.

One object of the present disclosure is to make the standardizedelectro-mechanical actuator a simple continuum of design choices amongswitched reluctance or brushless D.C. motors and star compound,multi-planet or eccentric single planet hypocyclic gear trains. Ideally,each choice can be considered as a plug-in substitute for the other withno other primary design changes.

Accordingly, certain embodiments of the present disclosure mayincorporate a single eccentric planet gear train in place of themulti-planet gear train used in FIGS. 1 and 2. The eccentric hypocyclicgear train incorporates a number of advantages, as described below.

In many embodiments, the actuators of the present disclosure incorporatea hypocyclic gear train, which may have a gear reduction ratio as highas 5000:1. These hypocyclic gear train assemblies may incorporaterelatively short circular arc gear teeth, with up to 5 or more teeth incontact at a time.

The unique design characteristics of the hypocyclic gear trains providereduced contact stresses by down to one tenth of known stresses, reducedbending stresses by down to one fifth of known stresses, lower slidingvelocity by down to one fifth of known velocities, reduced energy loss,and the potential for preloading the mesh as the tooth comes into itscentral position.

Each gear tooth can be profiled to be slightly preloaded as it comesinto its central position, in order to reduce the generation oflower-order harmonics and control backlash and lost motion. Thispreloading can be accomplished through the introduction of a slightinterference between that tooth and the mating teeth as that tooth comesinto its central position. In certain embodiments, a cavity may beintroduced within each tooth in order to tailor the stiffness of theteeth and reduce closing noise. In one embodiment, for example, aportion of the required compliance may derive from a partially compliantbearing between the driving eccentric and the wobble gear.

Circular arc tooth profile gear trains exhibit a reduced degree of wearand noise, smooth and gradual load transfer among the teeth, and areduced or eliminated necessity for critical tolerances, as circular arcteeth do not require the critical tolerances generally associated withinvolute teeth. A circular arc tooth profile can also exhibit increasedstrength, as clearances for external involute teeth are not required.Finally, in certain embodiments, a reduction in the sliding velocitybetween the mating gear teeth reduces the frictional losses within themechanism.

FIG. 3 depicts a cutaway isometric of a rotary actuator 90 incorporatingan eccentric hypocyclic gear train. Rotary actuator 90 incorporates acentral stationary shaft 110 holding support bearings 112 that supportthe rotating motor armature 108 that drives the eccentric 106. Supportbearings 104 on the eccentric 106 drive the wobble cylinder, whichcontains the planetary gears 100 and 102 that mesh with the bull gear 94and sun gear 96 separated by the principal cross-roller bearing orsimilar large diameter bearing 98.

Bull gear 94 is attached directly to the actuator shell 92 of rotaryactuator 90 while sun gear 96 is attached directly to the outputattachment plate 118. The motor armature 116 is also held stationary bythe actuator shell 92. End plate screws (not shown) assist in making theassembly rather direct, holding the stationary shaft 114 for supportbearings 112.

Bearing 112 in the output attachment plate 118 supports the end of thestationary shaft 114. Seal 120 separates the output attachment plate 118from the shell 302 and protects the cross-roller bearing 98 from theelements. This design incorporates an additional bearing 110 to supportthe motion and force variation on the eccentric 106.

Rotary actuator 90 is notable for its inherent simplicity. The motorfield 116 and rotating motor armature 108, eccentric 106, planetarygears 100 and 102, bull and sun gears 94 and 96, respectively, and theprincipal roller bearing 98 are the primary components of rotaryactuator 90. Secondary components include bearings 116, 118, and 112.The remaining components are stationary, machined components.

Even though rotary actuator 300 is able to provide very high powerdensity in a very small package, it can be adapted to a wide range ofapplication requirements by means of minor design changes, such asnumbers of gear teeth, motor winding characteristics and current andvoltage levels, as examples. The inherent simplicity and versatility ofrotary actuator 90 enables mass production of most of the subcomponents,thereby providing economies of scale and attendant cost reductions. Thecharacteristics of a particular embodiment of rotary actuator 90 may bescaled to one of a number of pre-selected standardized sizes, in orderto provide an “off-the-shelf” solution to the system designer. In oneexample of a standardized set of such actuators, sixteen separatestandardized scaled actuators can be provided to meet a wide range ofdesign applications. A set of actuators of the type shown in FIG. 3 maybe constructed according to standard sizes. As one example, a set ofseventeen actuator sizes spanning from 0.25″ diameter up to 90″ indiameter could support the construction of a large population ofmachines, rapidly assembled and made operational as needed.

Simplicity not only brings with it lower cost, it also results incomponents that are forgiving in their design, manufacture andoperation. In particular, rotary actuator 90 should be relativelyinsensitive to rather large variations in temperature.

The use of a hypocyclic gear train wherein up to approximately five ormore gear teeth can be in contact at a given time brings with it theability to carry very heavy loads, eliminate backlash, minimize lostmotion and resist high levels of shock with relatively modest levels ofgear tooth stress, thereby providing both high endurance and reducedwear.

The number of design parameters is rather low. They are, to a greatextent, independent choices, and each has clear and explicit meaning tothe designer. Hence, not only is rotary actuator 90 exceptional inperformance in terms of weight, volume, endurance, output inertia, andpower density, it is easily understood by most designers, helping toassure its acceptance in the design community.

As described above, the eccentric offset e within the hypocyclic geartrain is driven by an electric prime mover and supported by a bearing ona stationary shaft. Given N₁, N₂ to be the gear tooth numbers for thebull and sun gears, respectively, and N₁′, N₂′ those associated meshinggears on the wobble planet, then the total gear train ratio is givensimply by r=(N₁ ¹N₂)/(N₁ ¹N₂−N₁N₂ ¹). Note that the larger is theeccentric the greater must be the balancing mass for the wobble gearand, therefore, the overall weight of the actuator rises accordingly.

The ratio can range from 10-to-1 up to 5000-to-1, the higher ratiosdepending on the choice of gear tooth geometry that can be designed forhigh load capacity, low noise, high precision, or low cost depending onthe application. In certain embodiments, the appropriate ratio can beattained using meshing gears wherein the number of teeth between the twovaries by a single tooth. Note also that that the pressure angle may bereduced to at least 7 degrees, thereby reducing sliding velocity andinternal forces.

In connection with the hypocyclic gear train shown in FIG. 3, the wobblegears 100 and 102 are disposed side-by-side. This construction has atendency to improve rigidity. For lower gear train ratios, the diameterof wobble gear 100 may differ by as much as 30% or more from thediameter of wobble gear 102. In such a case, wobble gears 100 and 102may be disposed with one inside the other, so that all gear meshes occurin a single plane.

Not only can the hypocyclic gear train be directly plugged into any ofthe epicyclic designs, its key design parameters are always visible tothe designer, thereby removing the aura of black magic in this area ofdesign. Since the planet gear wobbles, it must be balanced by acounterweight. In many embodiments, the mass of the counterweightrequired is small relative to the mass of the planet gear itself

FIG. 4, depicts a cutaway isometric view of a rotary actuator 130 inaccordance with a fourth embodiment of the present disclosure. Therotary actuator 130 incorporates a central stationary shaft 156 holdingsupport bearings 155 that support the rotating motor armature 148 thatdrives the eccentric 157. Support bearings 144 on the eccentric 147drive the wobble cylinder, which contains the planetary gears 140 and142 that mesh with the bull gear 134 and sun gear 136 separated by theprincipal cross-roller bearing or similar large diameter bearing 138.

Rotary actuator 130 employs a pancake configuration that incorporates anSRM prime mover 150 to produce a high torque/low speed rotary actuator130.

Bearing 158 in the output attachment plate 160 supports the end of thestationary shaft 156. Seal 164 separates the output attachment plate 160from the shell 132 and protects the cross-roller bearing 138 from theelements.

FIG. 5 depicts a cutaway isometric view of a rotary actuator 170 inaccordance with certain embodiments of the present disclosure. Thisgeometrically different format for a hypocyclic actuator concept isshown in FIG. 5 and generally designated 170. As seen in FIG. 5, thebull gear 174 and stator 194 of actuator 170 are rigidly connected tothe outer shell 172 and closed at the end by end plate 184.

Armature 190 contains wobble plate gears 180 and 182, which mesh withbull gear 174 and sun gear 176. Sun gear 176 is separated from bull gear174 by the principal cross-roller bearing or similar large diameterbearing 178, which also may function as the principal bearing for thejoint of the machine into which rotary actuator 170 is incorporated.

Rotary actuator 170 further incorporates bearings 198 and 199 to preloadthe mesh of the wobble plate gears 180 and 182, so as to ensure thatthey do not separate and to reduce vibration and the effect of wear.

Bearings 198 and 199 are centered on a second eccentric offset of e,180.degree. out of phase with the wobble armature eccentric 186.Bearings 198 and 199 roll on cylindrical surfaces machined into the endplate 184 and output plate 192, both of which are concentric with thecenterline of the rotary actuator 170.

The high torque, low output velocity rotary actuator 170 shown in FIG. 5is a combination of a hypocyclic switched reluctance motor, which maygenerate up to five times higher torque than a standard switchedreluctance motor, and a hypocyclic gear train, which may have up to fivetimes higher load capacity than a similar epicyclic gear train.Accordingly, rotary actuator 500 can be said to have, in certainembodiments, an enhanced performance envelope up to 25 times higher thanprior designs.

This overall performance enhancement factor of 25 is achieved in rotaryactuator 170 with five basic parts, the removal of five additionalancillary bearings and few, if any, components incorporating dimensionshaving any critical tolerances.

In rotary actuator 170, the wobble motor armature 190 is incorporatedinto the same part as the wobble plate gear pair 180 and 182. Rotaryactuator 170 incorporates a number of distinct advantages over priordesigns, including:

The need for only one principal cross-roller bearing 178 and twoancillary bearings 198 and 199;

Simplified controller technology owing to the fact that each stator 194pole is switched on and off only once in a wave as the wobble motorarmature 190 walks through an angle of 360 degrees×e (where e is theeccentricity of the wobble configuration) during each electrical cycle.

The result of the above is a form of magnetic gearing where the electriccycle occurs at an angular velocity rate of l/e relative to therotational velocity of the wobble motor armature 190. Given an angularvelocity of the electrical field and the wobble speed w_(f)=w_(e)=6667with e=0.015, for example, the output attachment plate 192 would rotateat 100 RPM and the output velocity, w_(o), would equal 1 RPM given agear reduction ratio of 100. Because of this electrical wave, torqueripple is virtually non-existent. Also, given a value of e=0.015, abalancing mass at r=30e means that only 1/900, or 0.111%, of the mass ofwobble motor armature 190 needs to be removed to perfectly balancewobble motor armature 190. The attributes of actuator 170 are such thatcertain variations of this design may be employed effectively as aback-drivable generator to produce energy from a mechanical powersource, such as a wind turbine.

For at least the embodiments shown in FIGS. 3-5, the followingadditional specific comments apply. In certain embodiments, the gearteeth are circular arch teeth in order to enhance smoothness, reducenoise from gear tooth impact and reduce the contact Hertzian stress. Inother embodiments, triangular gear teeth may better satisfy theapplication requirements. In other embodiments, specialized gear toothgeometry may be included where unique application requirements must bemet.

Wiring may be disposed entirely in the stationary stator as part of theouter shell and bull gear. The force path through the actuator is short.Armatures may be solid or laminated metal. Few, if any, criticaldimensions are required, thereby reducing the influence of manufacturingtolerances and temperature variations on performance. The use of shortgear teeth reduces bending stresses and friction losses. The meshing ofup to thirty teeth picks up and releases the load slowly to reducenoise.

For at least the embodiments shown in FIGS. 3-5, each gear tooth can beprofiled to be slightly preloaded as it comes into its central position,in order to reduce the generation of lower-order harmonics and controlbacklash and lost motion. This preloading can be accomplished throughthe introduction of a slight interference between that tooth and themating teeth as that tooth comes into its central position. In certainembodiments, a cavity may be introduced within each wobble gear tooth inorder to tailor the stiffness of the teeth and reduce closing noise. Aspreviously mentioned, a portion of the required compliance may derivefrom a partially compliant bearing between the driving eccentric and thewobble gear.

Certain applications may require a rugged rotary actuator with a stiffoutput gear train of high reduction ratio in a compact configuration.Depending on the specifics, such an actuator may be driven either by acylindrical brushless D.C. Motor (DCM) or a pancake switched reluctancemotor (SRM) prime mover. FIGS. 6 and 7, respectively, are cutawayisometric views of these alternate embodiments.

Rotary actuator 200 of FIG. 6 has a “coffee can” profile, while rotaryactuator 250 of FIG. 7 has the shape of a circular pancake disk. Rotaryactuator 200 is designed for use in robotics, while rotary actuator 200is useful in confined spaces between two walls. Both rotary actuators200 and 250 are capable of producing relatively high torque atrelatively low speeds. All other things being equal, rotary actuator 200will generally have a higher maximum speed than rotary actuator 250 anda somewhat lower maximum torque.

FIG. 6 is a cutaway isometric view of a rotary actuator 200 with thefirst stage of the epicyclic gear train 230 inside the magnet cylinder218 of the relatively high speed D.C. motor field 228. The planets 232and 234 ride on bearings 236 in a planet cage 238 attached to the magnetcylinder 218, which, in turn, rides on bearings 660. This embodiment isideal for use in dexterous machines.

Planets 232 and 234 may form a Ferguson paradox configuration drivingmoving external sun gear 244 and stationary external bull gear 242attached to the central shaft 222 of rotary actuator 200. Central shaft222 is attached to the outer shell 202 using machine bolts (not shown).

In certain embodiments, the first stage may be designed to reduce itsinertia, as it experiences higher speeds and lower torque. Planet gears232 and 234 may be made relatively narrow and still carry the necessaryload. The specific design parameters of these planet gears 232 and 234are dictated by the application.

There will be a trade off between the size of the motor components (themagnet cylinder 218 and the field 228) and the outer diameter of thefirst stage epicyclic gear train 230. The smaller the internal diameterof magnetic cylinder 218 and field 228, the larger the torque produced.The stationary central shaft 222 is long in this design and subject toflexure. It is, therefore, supported by bearing 230.

Sun gear 244 is rigidly connected to the driving cage 216 of the secondstage epicyclic gear train 246 riding on large needle bearings 220carrying planet gears 210 and 212 riding in bearings 214. These planetgears 210 and 212 mesh with stationary internal bull gear 204, which isattached to the outer shell 202, and internal sun gear 206 is attacheddirectly to the output attachment plate 226.

Seal 248 separates the attachment shell 202 from the plate 226. Externalsun gear 244 and its planet cage 238 support a bearing 249, which isheld in place by the outer shell 202. The shape of outer shell 202supporting bearing 249 not only strengthens the outer shell 202 but alsoimproves the rigidity of the central stationary shaft 222. Internal sungear 206 is rigidly attached to the output attachment plate 226, whichcontains bearing 224, to further strengthen the output structure ofrotary actuator 200.

The second stage 246 of the epicyclic gear train uses an internal bullgear 204 and sun gear 206. This arrangement conforms to the basicconfiguration of the structure, minimizing weight while at the same timemaking rotary actuator 200 particularly rugged and stiff

In the second stage 246, the velocities are lower so the concern forinertia goes down accordingly, but the regard for stiffness and loadcapacity go up. Hence, the size of the gear teeth in the second stage246 must meet the requirement for load as a first priority, withstiffness as a second priority. This may require, in certainapplications, the use of as many planets 210 and 212 as the geometrywill allow.

The principal bearing in this configuration is the cross-roller bearingor similar large diameter bearing 208. It separates internal bull gear204 and shell 202 from internal sun gear 206 and output attachment plate226. Cross-roller bearing 208 also performs the load bearing tasks forthe machine using this actuator. Because of the position of cross-rollerbearing 208, internal bull gear 204 can be made very stiff, as caninternal sun gear 206. For maximum stiffness and minimum deflectionunder load, the attachments to the neighboring links should be madeclose to cross-roller bearing 208.

FIG. 7 depicts, in a cutaway isometric view, an embodiment of a rotaryactuator 250 of the present disclosure configured for a relatively lowspeed pancake SRM, which produces relatively high torque. The bull gear254 is made especially strong and is rigidly attached to the attachmentshell 252 and supporting bearing 732 to the primary stationary shaft270, so as to further strengthen the output attachment plate 274 forthis design.

Magnet disk 268, in concert with field 276, directly drives the firststage planet cage 266 for planet gears 260 and 262, which are supportedin bearings 264. The first stage planet cage 268 must be carefullydesigned to accommodate the planet gears 260 and 262 while maintainingsufficient structural integrity.

The second stage planet cage 277 is driven by, and rigidly attached to,the first stage sun gear 289, which is supported by three bearings 278,280 and 282 in order to maximize its support. This support isincorporated to resist twisting moments generated by the second stageplanets 284 and 286 supported in bearings 287. The first stage sun gear289 and bull gear 288 are external gears. Bearing 272 supports the firststage planet cage 266 in the first stage sun gear 289, which drives thesecond stage planet cage 277.

The second stage sun gear 256 and bull gear 254 are internal gears. Thisarrangement serves to match the structural geometry of the rotaryactuator 250 so as to stiffen the structure. The second stage sun gear256 and bull gear 254 are separated by the principal cross-rollerbearing or similar large diameter bearing 258 which acts as theprincipal bearing in the gear train while also serving as the principalbearing of the joint into which the rotary actuator 250 is incorporated.In order to maximize rigidity, the attachments to the attachment shell252 and to the output attachment plate 274 should be placed close tocross-roller bearing or similar large diameter bearing 258.

Since the second stage bull gear 254 and sun gear 256 are relativelylarge in diameter, they are able to accommodate more second stageplanets 284 and 286 and larger gear teeth. Accordingly, second stageplanet gears 284 and 286 are shown to be relatively large as compared toplanet gears 260 and 262 in FIG. 7.

Because of the lower speeds encountered in the second stage gear train,concern for inertia is superseded by a concern for the load capacity ofthe gear teeth. This is also true, to a lesser extent, in the firststage of the gear train. The outer envelope of the first stage issmaller in diameter than the outer envelope of the second stage, whichis appropriate since it carries less load but operates with largerangular velocities.

FIG. 8 depicts a cutaway isometric view of a rotary actuator 290incorporating a quick-change attachment architecture in accordance withcertain embodiments of the present disclosure. Rotary actuator 290incorporates an actuator shell 292 containing a bull gear 294, and sungear 296, separated by a cross-roller bearing or similar large diameterbearing 298. Planet gears 300 and 302 mesh with bull gear 294 and sungear 296, respectfully.

As seen in FIG. 8, actuator 290 rigidly connects a first mechanical link304 to a second mechanical link 322. First mechanical link 304 isrigidly connected to actuator shell 292 by a first wedge clamp 308,while second mechanical link 306 is rigidly connected to outputattachment plate 318 by second wedge clamp 310. In one embodiment, eachof wedge clamps 308 and 310 takes the form of a pair of semi-circularwedge clamp halves tightened against actuator 290 by an external bandclamp. Other equivalent structures may, of course, be employed withoutdeparting from the spirit and scope of the present disclosure.

In the embodiment shown in FIG. 8, wedge clamps 308 and 310 aretightened by a pair of tensioning mechanisms 312 and 314. Depending onthe particular application, tensioning mechanisms 312 and 314 may beintegral to the wedge clamps 308 and 310, or they may be integral toseparate band clamps disposed around wedge clamps 308 and 310.

Each of wedge clamps 308 and 310 incorporates a pair of generally-conicinternal surfaces, together forming a groove about the internal surfaceof the wedge clamp 308 and 310. The internal profile of each of theseinternal surfaces conforms to a mating external surface on either theactuator 290 or one of the mechanical links 304 and 306. As thetensioning mechanisms 312 and 314 are tightened, the normal forcebetween the generally-conic internal surfaces and the mating externalsurfaces will draw the mating components together into a tight and rigidmechanical connection. In certain embodiments, the design of wedgeclamps 308 and 310 will conform to one of a standard set of sizes.Within each standard size, there may be two or more strength classes,similar to the types of classification employed for standardizedthreaded fasteners.

Mechanical links 304 and 306 are disposed closely adjacent to oneanother and to principal cross-roller bearing 298. With the attachmentof mechanical links 304 and 306 in this manner, closely adjacent to oneanother and to principal cross-roller bearing 298, it can be seen thatthe joint rigidly resists motion about five of the six degrees offreedom, with the remaining degree of freedom controlled by the primemover and gear train combination.

It can be seen that the “force path” through the rotary actuator 290 isextremely short, and passes through a combination of highly rigidmechanical structures and connections and associated rigid structures.This short force path and associated rigid structures enable the rotaryactuator 290 to serve as the rotary joint for the machine itself, ratherthan serving merely as a torque input device, as in prior designs.

It will be appreciated by those of skill in the art that, although thequick-change attachment structures of rotary actuator 290 are shown inconnection with a particular embodiment of the present disclosure, theattachment structures shown in FIG. 8 can be employed in connection withany of the embodiments described herein without departing from thespirit and scope of the present disclosure. Where simplicity is desired,simple bolt circles may prove adequate where accuracy and repeatabilityof the interface are not high priorities, or where a quick change of theactuator out of the system is not considered important to theapplication.

The structures shown and described in connection with FIG. 8 apply toall rotary actuators described herein. The geometry of a machine builtfrom the actuators described herein will be primarily dependent on themembers attached to the actuators rather than on the actuatorsthemselves. Depending on the application, the links may be parallel toone another, perpendicular to one another, or disposed at any generalspatial orientation to one another. The link geometry provides a machinedesigner with a great deal of freedom to design the system without thenecessity for customized components. The use of standardized componentscan, in many instances, reduce cost, owing to the availability of massproduction of both the actuators and the links connecting them. At thesame time, a high degree of generality and flexibility can be preservedfor the designer, even when using standardized components.

FIG. 9 shows a side view of a circular arc gear tooth mesh in accordancewith certain embodiments of the present disclosure. Specifically, itshows the sequence of motion, within a sun/bull gear mechanism 320, of asun gear tooth as it enters and exits its central position within thebody of the stationary bull gear 322.

The initial position of the sun gear tooth at time T0, prior toengagement with the bull gear 322 is designated 324. The centralposition of the sun gear tooth at time T1, some period of time aftertime T0, is designated 324′.

In certain embodiments, the geometry of mechanism 320 may be such that aslight interference is encountered as the sun gear tooth moves into thecentral position 324′. In such embodiments, the gear tooth stiffness andthe level of interference in the central position 324′ will determinethe forces generated by the elastic deformation of the bull gear 322 andthe top of the sun gear tooth. This interference will tend to reduce oreliminate any free motion in any of the bearings supporting the sungear. It can be seen in FIG. 9 that the sun gear tooth 324 shownincorporates a cavity in order to reduce its stiffness, as will bedescribed in more detail below in connection with FIGS. 10-12.

After time T1, at which point maximum interference and deformation, ifany, occur, the sun gear tooth 324′ will move out of engagement with thebull gear 322. The position of the sun gear tooth at a point in time T2after time T1 is designated 324″.

Examples of gear tooth geometry useful in connection with gear mechanism320 and similar gear mechanism are shown in FIGS. 10-12.

FIG. 10 depicts a side view of a circular arc gear tooth 330 having abody 332, a first flank 334, a second flank 336, and a circular cavity338 disposed at the top of the body 332. The position and diameter ofcavity 338 will be determined by the requirements of a particularapplication. In general, the stiffness at the peak of gear tooth 330will be reduced as the diameter of the cavity 338 is increased or itscentral axis is moved closer to the peak of gear tooth 330. Reducing thediameter of the cavity 338 or moving it further down into the body 332will have the opposite effect, tending to stiffen the peak of gear tooth332.

FIG. 11 depicts a side view of a circular arc gear tooth 340 having abody 342, a first flank 344, a second flank 346, and a circular cavity348 disposed at the top of the body 342. Gear tooth 350 furtherincorporates a slot 350 at the top of circular cavity 348, so as toreduce the rigidity of the top of the body 342 of gear tooth 350.

FIG. 12 depicts a side view of a circular arc gear tooth 360 having abody 362, a first flank 364, a second flank 366, and a cavity 368disposed at the top of the body 362. Cavity 368 is composed of twocircular cavities 370 and 372, which overlap in the center of gear tooth360. This design preserves the local stiffness at the top of the geartooth 360.

In essence, therefore, the disclosed subject matter provides aself-contained integrated actuator that combines a prime mover and geartrain for yielding a compact rotary actuation torque within a largersystem. The self-contained integrated actuator includes a cross-rollerbearing for operating as a structural joint in the larger system ofwhich the self-contained integrated actuator is a part. An outerattachment shell for rigidly interfacing the larger system andcontaining a motor stator and an internal bull gear, the motor statorfor generating a controllable electromagnetic field, and the internalbull gear for interfacing the cross-roller bearing and providingstiffness for the self-contained integrated actuator, and furthercomprising a plurality of internal gear teeth. An output attachmentplate contains an internal ring gear and supporting a plurality of driveshaft bearings. The internal ring gear rigidly interfaces the outputattachment plate and further includes a plurality of internal gearteeth, the cross-roller bearing further for positioning the outputattachment plate within the outer attachment shell. A drive shaft holdsa prime mover rotor and an eccentric and associating with the outputattachment plate via the plurality of drive shaft bearings, the primemover rotor rotates in response to the controllable electromagneticfield and the eccentric.

A gear train associates with the eccentric and includes the meshinggear, wherein the meshing gear further comprises a plurality of externalgear teeth, the external gear teeth includes circular arc surfaces thatmesh with the plurality of internal gear teeth of the internal bull gearand the plurality of internal gear teeth of the internal ring gear, thegear train walk a minimal number of the plurality of external gear teethfor each rotation of the prime mover rotor. The cross-roller bearing,the outer attachment shell, the drive shaft, and the meshing gear traincooperate to provide a self-contained integrated actuation torquetransmitting force from the prime mover through the gear train along ashortest-possible transmission path.

The cross-roller bearing, the outer attachment shell, the drive shaft,the meshing gear train, and the prime mover associate in the form of apancake-shaped self-contained integrated actuator having a diameter atleast equal to approximately the length of the cylindricalself-contained integrated actuator. The self-contained integratedactuator may, for example, either associate the cross-roller bearing,the outer attachment shell, the drive shaft, the meshing gear train, andthe prime mover associate in the form of a cylindrical self-containedintegrated actuator having a height at least equal to approximately thediameter of the cylindrical self-contained integrated actuator.

The self-contained integrated actuator may be configured so that theinternal bull gear, the internal ring gear, and the meshing gear orwobble gear form a hypocyclic gear train or include a planet gear forforming an epicyclic gear train. Alternative, the meshing gear may be afixed axis gear for forming a star-compound gear train.

The self-contained integrated actuator may be formed to be of a sizespecified by a predetermined set of standardized dimensions. Thestandardized dimensions may accord with predetermined certificationrequirements for use of the self-contained integrated actuator. Also,the predetermined certification requirements may define a minimal numberof standardized dimensions for a maximal variety of uses of theself-contained integrated actuator. The internal gear teeth of theinternal bull gear, the internal gear teeth of the internal ring gear,and the external gear teeth of the meshing gear may associate with atleast a 75-to-1 gear reduction ration.

The self-contained integrated actuator may further include aquick-change interface for rapid replacement of the self-containedintegrated actuator. The quick-change interface accommodates rapidreplacement of the self-contained integrated actuator, wherein thequick-change interface has the form of a predetermined subset of aplurality of predetermined quick-change interface forms. In addition,the quick-change interface may provide a resistive force in up to sixdirections, at least a subset of the up to six directions demonstratinga predetermined degree of structural stiffness. As such, thequick-change interface is positioned immediately proximate to thecross-roller bearing for providing a minimal force path from the primemover through the gear train to the quick-change interface.

The prime mover may operate at a maximal power density frequency rangingup to approximately 30,000 revolutions per minute. The gear train mayoperate at reduction rates ranging between approximately 75-to-1 and5,000-to-1 and further include a hypocyclic gear train for operating atreduction rates ranging between approximately 75-to-1 and 5,000-to-1.Alternatively, the gear train may further include a star-compound geartrain for operating at reduction rates ranging between approximately5-to-1 and 25-to-1. The gear train may further include an epicyclic geartrain for operating at reduction rates ranging between approximately5-to-1 and 75-to-1.

The output attachment plate provides an output torque for maximal torquedensity ranging from approximately 400 to 600 inch-pounds per pound. Thegear train further includes a two-stage configuration, the two-stageconfiguration further includes a star compound gear train and anepicyclic gear train configuration. Alternatively, the gear trainfurther may include a two-stage configuration, the two-stageconfiguration further includes an epicyclic gear train and an epicyclicgear train. The internal gear teeth of the internal bull gear, theinternal gear teeth of the internal ring gear, and the external gearteeth of the meshing gear mesh with a pressure angle of less thanapproximately nine degrees.

In its simplest form, the self-contained integrated actuator consistsessentially of the cross-roller bearing, the outer attachment shell, thedrive shaft, the single-stage meshing gear, and the prime mover fortolerance insensitivity, temperature insensitivity, and increasedendurance.

A preloading force applied to the internal bull gear, the internal ringgear, and the wobble gear substantially eliminate backlash in theoperation of the self-contained integrated actuator. The preloadingforce may be derived from interfacing the external gear teeth, theplurality of internal gear teeth of the internal bull gear and theplurality of internal gear teeth of the internal ring gear in a cuspmotion perpendicular to the meshing gear. In addition, the disclosedsubject matter provides for load sharing among a plurality of theinternal teeth of the internal bull gear, a plurality of the internalteeth of the internal ring gear, and a plurality of the external teethof the meshing gear, thereby substantially eliminating lost motion inthe operation of the self-contained integrated actuator.

Concave-convex tooth contact among a internal teeth of the internal bullgear, internal teeth of the internal ring gear, and external teeth ofthe meshing gear for substantially eliminate lost motion in theoperation of the self-contained integrated actuator. In particular, theinternal teeth of the internal bull gear, the internal teeth of theinternal ring gear, and the external teeth of the meshing gear form aforce distribution characteristic in the form of a symmetric parabola.

A plurality of wedge clamps on predetermined sides of the self-containedintegrated actuator attach neighboring links to the self-containedintegrated actuator. Each of the wedge clamps includes two semi-circularportions for resisting primary opening forces applying to each of thewedge clamps. A band clamp secures each of the wedge clamps into apredetermined position. A plurality of attachments are immediatelyproximate to the cross-roller bearing.

The self-contained integrated actuator further includes a plurality ofoperational sensors for sensing operational characteristics of elementsfrom the group consisting essentially of the cross-roller bearing, theouter attachment shell, the drive shaft, the meshing gear train, and theprime mover. The operational sensors may sense operationalcharacteristics of elements from the group consisting essentially of thecross-roller bearing, the outer attachment shell, the drive shaft, themeshing gear train, and the prime mover for maintaining an optimaloperational envelope for the self-contained integrated actuator, as wellas for maintaining an optimal operational maintenance schedule for theself-contained integrated actuator.

In the embodiments described above, the tooth ends may need moreductility than the remainder of the tooth surface, which shouldgenerally be hardened. In certain embodiments, the cavity or cavitiesmay be drilled and/or slotted before hardening. The tooth surface maythen be hardened. The tooth tips may be annealed locally to improve thefatigue resistance at the deforming part of the tooth. This annealingmay, in certain embodiments, be performed by a laser.

To be certain, the present disclosure provides an integratedstandardized rotary actuator incorporating a prime mover, a gear train,and a rotary machine joint in a single package. These elements areintegrated into a single self-contained module that is easily scalableto meet a wide variety of application demands. The rotary actuator mayincorporate as few as five principal parts fitted with a minimum ofcritical tolerances, resulting in a system that is substantiallyinsensitive to tolerance and temperature variations. Applications forthe various embodiments here described include a driving force foressentially any machine or apparatus that moves. Transporters infactories, food machinery, and packaging, conveyor, and handling systemmake ideal applications for the disclosed subject matter. Emergingapplications may include space-based, remotely controlled systems (FIG.13) entertainment systems, educational robots, surgical systems, glovebox systems, farm machinery, construction machinery, buses (hybrids),trucks (hybrids) (FIG. 14), elevators, and wind turbines (FIG. 15).Military applications include ships and submarines, aircraft and UAVs,anti-terrorism robots (FIG. 16), AND tanks (including 20 ton vehicles)(FIG. 17). Industrial applications may also include manufacturing celland robotics applications (FIG. 18). Still further applications mayinclude automobiles and humanoid or prosthesis applications. All suchapplications as these may provide a significantly improved overallsystem capable of making novel and beneficial use of the presentlydisclosed self-contained, integrated gear train and prime moveractuation module.

The electric prime mover 380 drives the pinion gear 382 of the reducer(see FIGS. 19 and 20). The pinion gear 382 drives three (or more) stargears 384 with stationary shaft bearings. Each shaft contains eccentric386 (all the same size) with value e, which are completely in parallelwith each other (in phase). Eccentrics 386 can be thought of as threeparallel/in-phase driven crankshafts. Each eccentric drives the paralleleccentric gear 388 (wobble gear, or similar functioning component)through a bearing. The parallel eccentric gear 388 exhibits a circularmotion (without rotation) which in itself is unbalanced. To counteractthis imbalance, the eccentrics 386 have another eccentric with magnitudek×e which drives a PE balance mass 390. The parallel eccentric mass 390must create an opposite inertia force to balance that of the paralleleccentric gear 388 (which includes the other unbalanced massesassociated with the eccentrics and crankshaft bearings). Hence, theweight of the parallel eccentric mass 390 can be reduced according tothe magnitude of the eccentric ratio k.

The parallel eccentric gear 388 contains an external toothed gear on itsperiphery. It meshes with the internal teeth of the output ring gear392. The relative motion between the parallel eccentric gear 388 and theoutput ring gear 392 is that the parallel eccentric gear 388 rollsinside the output ring gear 392. This relative motion is calledhypo-cycloidal motion.

A relatively small pinion can drive the concentric star gears such thattheir ratio can be 1-to-1 up to 4-to-1 in size. Given a ratio of 1-to-1requires that the pinion increase in diameter such that diametric circleof the eccentric cranks be as large as possible. In this design, most ofthe compliance at the output will be due to the radial stiffness of theeccentric crank shaft bearings. Hence, larger bearings and larger piniondiameter become necessary to increase the parallel eccentric geartrain's (PEGT) output stiffness. Given a 1-to-1 star ratio, the totalratio of the train is that of the mesh of the parallel eccentric gearwith the ring gear. Given 52 teeth in the output ring gear and 49 teethin the parallel eccentric gear should give a ratio of 17.33-to-1 (i.e.,the number of teeth divided by the tooth difference). It is entirelypossible for this to be 52-to-1 if there is only one tooth difference.Or if the star compound ratio goes to 4-to-1, then 4×17.33=69.3-to-1becomes the total reduction ratio. Hence, this parallel eccentric geartrain provides a new level of freedom for the designer to optimize thegear train in the reduction range from 25-to-1 up to say 200-to-1.Larger values for the star compound ratio reduce the effective mass ofthe parallel eccentric gear train. Smaller tooth differences between theparallel eccentric gear and the output ring gear (enables smallereccentrics) allows the use of more and shorter teeth enhancing loadcapacity and creating greater smoothness in the mesh upload/downloadcycle.

The parallel eccentric gear trains employ a recommended principalbearings (a cross-roller bearing) 394 between the shell backbonestructure 396 and the output ring gear 392. Secondary support bearing398 is positioned between the output ring gear 392 and the gear cages400. The principal bearing 394 and support bearing 398 may be similar(in size) tapered roller or 4-point ball bearings in order to reducecost.

A second version of the parallel eccentric gear train is shown in FIGS.21 and 22. Here, the wobble gear 410 in FIG. 19 is split into two gears(both mesh with the same output ring gear) which are driven bycrankshaft eccentrics 412 180° out of phase, which automaticallybalances the masses. It has the same load capacity in the tooth mesh(the same tooth width) and the tooth load is divided in two locations onthe output ring gear 414. It is likely that the crankshaft bearings willindividually be 50% of the width of that in FIGS. 19 and 20 (to providefor the same overall size and weight), such that the load capacity willbe similar to that for the parallel eccentric gear train in FIGS. 19 and20. Its stiffness, however, will go up by approximately 50%.

The parallel eccentric gear trains employ a recommended principalbearings (a cross-roller bearing) 416 between the shell backbonestructure 418 and the output ring gear 414. A secondary support bearingis positioned between the output ring gear 414 and the gear cages 420.The principal bearing 416 and support bearing may be similar (in size)tapered roller or 4-point ball bearings in order to reduce cost.

A third version of this parallel eccentric gear train is laid out inFIGS. 23 and 24. In this case, a single parallel eccentric gear 430 iscombined with a central (extra) crankshaft along the gear train's centerline. This requires that the central eccentric, input pinion 432, andthe three concentric eccentrics (crankshafts) 434 all rotate in unison(parallel) and in the same direction. This necessitates the use of idlergears 436 between the input pinion 432 and the star gears 438 on theconcentric eccentrics (crankshaft) 434. This is called the PEx geararrangement (x for extra crankshaft), and will refer to the gear trainas PE×GT. The central crankshaft, input pinion 432, ensures an equalload distribution among the concentric eccentrics (crankshafts) 434.Relative to the parallel eccentric gear train in FIG. 19, this PExdesign should be approximately 2.4× stiffer and provide for 3× largerload carrying capacity. (Note that these results depend on the samecriteria for the crankshaft bearing selection in both designs).

The parallel eccentric gear trains employ a recommended principalbearings (a cross-roller bearing) 440 between the shell backbonestructure 442 and the output ring gear 444. Secondary support bearing446 is positioned between the output ring gear 444 and the gear cages448. The principal bearing 440 and support bearing 446 may be similar(in size) tapered roller or 4-point ball bearings in order to reducecost.

A fourth vision of the parallel eccentric gear train is provided inFIGS. 25 and 26. The principal issue being addressed is the need to notonly have the strengthening central shaft but also a pair of (dual)eccentric gears 460 operating 180° out of phase to better carry theexternal load at the three concentric eccentrics (crankshafts) 462. Inthe parallel eccentric design, the critical load elements are thebearings on the three concentric crankshafts. This dual eccentricconfiguration essentially doubles their number in a very compactarrangement. The input pinion 464 drives three concentric amplifiergears 466 in a star compound arrangement which replace the role of theidler gears in the PE×GT of FIGS. 23 and 24. The amplifier gears 466then drive star gears 468 attached to each of the three concentriceccentrics (crankshafts) 462 to make a very compact input structure forthe dual eccentric gears 460 with the “extra” central crankshaft 470.The extra crankshaft 470 enables the external load on the gear train tobe evenly distributed among the three concentric eccentrics(crankshafts) 462. The goal is to lighten all gears by using thin webstructures between their support bearings and their external teeth.Also, the maximum size bearing is intended for all crankshaft bearings.

The parallel eccentric gear trains employ a recommended principalbearings (a cross-roller bearing) 472 between the shell backbonestructure 474 and the output ring gear 476. Secondary support bearing478 is positioned between the output ring gear 476 and the gear cages480. The principal bearing 472 and support bearing 478 may be similar(in size) tapered roller or 4-point ball bearings in order to reducecost.

As noted above, rotary actuators may be made in accordance with theteachings herein whose construction is based on shortest force pathprinciples. Such a construction is depicted in FIG. 27, whichillustrates the envelop 89 through which the shortest force path for therotary actuator 50 of FIG. 2 passes. As seen therein, the principalbearing 58 (which is preferably a cross-roller bearing) has a firstsurface which is attached to the output attachment plate 84, and asecond surface which is attached to the outer attachment shell 52.Hence, force passes from the outer attachment shell 52 through theprincipal bearing 58 and through the output attachment plate 84.

In a preferred embodiment of a rotary actuator made in accordance withthe foregoing principles, the output attachment plate 84, outerattachment shell 52 and principal bearing 58 are arranged such that theshortest force path passes through the output attachment plate 84, theprincipal bearing 58 and the outer attachment shell 52 (preferably inthat order in going from the output attachment plate 84 to the outerattachment shell 52) along an axis perpendicular to the outputattachment plate 84. The envelop 89 through which the shortest forcepath for the rotary actuator 50 of FIG. 2 passes preferably includes aline, more preferably a cylinder (which is preferably a right cyclinderwhose axis is perpendicular to said output attachment plate 84), andmost preferably an annulus. Such an arrangement ensures that theshortest force path through these elements and/or through the rotaryactuator 50 is attained. The advantage of a shortest force path whichincludes an annular envelop may be appreciated by considering a singletooth gear mesh for external rotating gears and the locking effect ofthe two face gears (as in synchro mesh gear clutches), where all teethshare equally in the load (the ideal condition).

An arrangement of this type also enhances or optimizes stiffness in theactuator. For example, a bending component in an actuator typicallyreduces stiffness by three orders of magnitude (400×), while anytorsional component is typically 2500× less stiff than an equivalentlinear path.

The use of a cross-roller bearing in such a configuration also enhancesstiffness, since the roller bearing will always be the weakest componentin the linear force path. In this respect, it is to be noted that singlepoint race contact ball elements are the least stiff, while double pointrace contact ball elements are roughly 50% more stiff than acorresponding single point race contact ball element. However, across-roller bearing has a linear contact with the race which is roughly6× stiffer than a 2-point race contact ball element, and 9× as stiff asa single-point race contact ball element. In this respect, it is to benoted that radial and bending loads effectively load only a smallportion of the rolling elements, with a gradual drop-off of contactforce from the principle contact force or center. Hence, an important(and frequently dominant) objective is to utilize the stiffest bearingso as to make maximum linear contacts, thus maximizing durability andpreventing non-compatible tolerances or deformations among multiplebearings that are doing the same task. It is found that actuatordurability is largely determined by the wear history of the bearings,since the gear mesh is increasingly better able to carry out itsfunction by continuously improving designs.

Various embodiments of actuators made in accordance with the foregoingprinciples may be constructed in which the shortest force pathconfiguration may be expressed in various ways which may vary from oneembodiment or application to another. For example, if the envelop 89through which the shortest force path for the rotary actuator 50 of FIG.2 passes is a cyclinder, then in some embodiments, it is preferred thatthe locus formed by the intersection of the cylinder and a first majorsurface (which may be the exterior surface) of the output attachmentplate 84 is a circle.

Symmetry, especially axial or rotational symmetry, is a desirablefeature in the actuators described herein. Hence, in a preferredembodiment of an actuator made in accordance with the teachings herein,the output attachment plate rotates about a central axis, and the line,cylinder or annulus containing the shortest force path is preferablyparallel to this axis of rotation. Even more preferably, the rotaryactuator, the output attachment plate, the principal bearing, and/or theoutput attachment shell, are axially or rotationally symmetric about theaxis of rotation (which is preferably the central axis of the actuator).Symmetry is desirable in these actuators, because deviations fromsymmetry may distort the symmetric rolling element force distribution,thus causing some elements to carry a correspondingly higher load (thusadversely affecting durability and stiffness and thereby reducingoverall load carrying capacity per unit weight).

It will be appreciated that, in the various embodiments of rotaryactuators described herein, various considerations or factors may needto be weighed against each other in designing a rotary actuator for agiven application. Hence, the desire to minimize weight, volume, cost orthe effects of tolerance errors, and/or the desire to maximizestiffness, make the best use of compatible materials, match coefficientsof thermal expansion, or to force fit assemblies, may need to bebalanced against each other to arrive at a suitable actuator solution.In many applications, cost may be the overriding factor. Hence, for lowcost actuator solutions, lower constraints on the design objectives mayoccur, while on higher end solutions, tighter constraints or betterdesign criteria may be employed.

As previously noted, the rotary actuators disclosed herein may beutilized as structural joints for the devices they are incorporatedinto. In particular, the rotary actuators disclosed herein preferablycombine the traditional functionality of rotary actuators as torquegenerating devices with the ability of the actuator to act as astructural joint. These aspects may be incorporated into plug-and-playopen architecture systems to provide a much more self-contained minimumset of actuators for a maximum range of solutions in the openarchitecture for a given domain (e.g., motorized vehicle applications).This approach may be effectively utilized to substantially reduceweight, volume and cost in the resulting solution set.

Although preferred embodiments of the disclosure have been described indetail, it will be understood by those skilled in the art that variousmodifications can be made therein without departing from the spirit andscope of the disclosure as set forth in the appended claims.

1. A rotary actuator, comprising: an output attachment plate; an outerattachment shell; and a principal bearing having a first surface whichis attached to said output attachment plate, and a second surface whichis attached to said outer attachment shell.
 2. The rotary actuator ofclaim 1, wherein said output attachment plate has a first major surface,and wherein said output attachment plate, said outer attachment shelland said principal bearing are arranged such that a first line existswhich is perpendicular to said first major surface of said outputattachment plate and which passes through said output attachment plate,said principal bearing and said outer attachment shell.
 3. The rotaryactuator of claim 2, wherein said first line passes through said outputattachment plate, said principal bearing and said outer attachmentshell, in that order.
 4. The rotary actuator of claim 3, wherein saidoutput attachment plate rotates about a central axis, and wherein saidfirst line is parallel to said central axis.
 5. The rotary actuator ofclaim 4, wherein said principal bearing is a crossed roller bearing. 6.The rotary actuator of claim 4, wherein said output plate is axiallysymmetric about said central axis.
 7. The rotary actuator of claim 6,wherein said outer attachment shell is axially symmetric about saidcentral axis.
 8. The rotary actuator of claim 7, wherein said principalbearing has rotational symmetry about said central axis.
 9. The rotaryactuator of claim 1, wherein said output attachment plate, said outerattachment shell and said principal bearing are arranged such that afirst cylinder exists which passes through said output attachment plate,said principal bearing and said outer attachment shell.
 10. The rotaryactuator of claim 9, wherein said first cylinder is a right cylinder.11. The rotary actuator of claim 10, wherein said first cylinder has anaxis which is perpendicular to said output attachment plate.
 12. Therotary actuator of claim 11, wherein said output attachment plate has afirst major surface, and wherein the locus formed by the intersection ofsaid first cylinder and said first major surface of said outputattachment plate is a first circle.
 13. The rotary actuator of claim 12,wherein said first major surface is a second circle, and wherein saidfirst and second circles are concentric.
 14. The rotary actuator ofclaim 13, wherein said first major surface is an exterior surface. 15.The rotary actuator of claim 12, wherein said output plate rotates abouta central axis, and wherein said output plate is axially symmetric aboutsaid central axis.
 16. The rotary actuator of claim 15, wherein saidouter attachment shell is axially symmetric about said central axis. 17.The rotary actuator of claim 16, wherein said principal bearing hasrotational symmetry about said central axis.
 18. The rotary actuator ofclaim 1, wherein said output attachment plate contains a first annulus,wherein said outer attachment shell contains a second annulus, andwherein said principal bearing contains a third annulus, and whereinsaid output attachment plate, said outer attachment shell and saidprincipal bearing are arranged such that said first, second and thirdannuli are concentric.
 19. The rotary actuator of claim 18, wherein acylinder exists whose surface intersects and is concentric with saidfirst, second and third annuli.
 20. The rotary actuator of claim 19,wherein said cylinder is a right cylinder.
 21. The rotary actuator ofclaim 18, wherein said output plate rotates about a central axis, andwherein said output plate is axially symmetric about said central axis.22. The rotary actuator of claim 21, wherein said outer attachment shellis axially symmetric about said central axis.
 23. The rotary actuator ofclaim 22, wherein said principal bearing has rotational symmetry aboutsaid central axis.
 24. The rotary actuator of claim 1, wherein saidoutput attachment plate, said outer attachment shell and said principalbearing are arranged such that a force applied in a first directionnormal to said output attachment plate is transmitted through saidprincipal bearing and said outer attachment shell with essentially notransmission of force in a direction orthogonal to said first direction.25. The rotary actuator of claim 1, wherein said outer attachment plateand said outer attachment shell are rigid in a direction orthogonal to amajor surface of said outer attachment plate.
 26. The rotary actuator ofclaim 1, wherein said outer attachment plate and said outer attachmentshell are rigid in a direction orthogonal to a major external surface ofsaid outer attachment plate.
 27. A machine joint which includes therotary actuator of claim 1.